Moving crank shaft force point

ABSTRACT

Embodiments of the present invention relate to lifting and lowering loads more efficiently and also more economically than known systems. This invention is the continuous movement by mechanical means of a moving crank shaft force point to a desired advantageous position at a desired advantageous moment to achieve improved geometry resulting in a desired low torque factor for a reduced net torque when lifting or lowering an unbalanced load with a beam with a fulcrum and connected to a load and an effort. In one embodiment, a walking beam well pumping unit, the lifting and lowering of the well load can be caused by the reciprocating motion of a beam tipping on a fulcrum and with an effort force point continuously substantially vertically disposed with the moving crank shaft force point. Potentially reduced net torque might allow longer life speed reducers, smaller speed reducers, and longer reciprocating vertical stroke length.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims priority to and the benefit of the filing of U.S. Provisional Patent Application No. 62/411,556, filed on Oct. 22, 2016; U.S. Provisional Patent Application No. 62/403,165, filed on Oct. 2, 2016; U.S. Provisional Patent Application No. 62/421,410, filed on Nov. 14, 2016; U.S. Provisional Patent Application Ser. No. 62/426,337, filed on Nov. 25, 2016; U.S. Provisional Patent Application Ser. No. 62/535,846, filed on Jul. 22, 2017; U.S. Provisional Patent Application Ser. No. 62/535,945, filed on Jul. 23, 2017; and U.S. Non-Provisional patent application Ser. No. 15/719,964, filed on Sep. 29, 2017; U.S. Non-Provisional patent application Ser. No. 15/789,760, filed on Oct. 20, 2017; U.S. Non-Provisional patent application Ser. No. 15/810,083, filed on Nov. 12, 2017; U.S. Non-Provisional patent application Ser. No. 16/130,068, filed on Sep. 13, 2018; Non-Provisional patent application Ser. No. 16/709,679, filed on Dec. 10, 2019; Non-Provisional patent application Ser. No. 16/709,815, filed on Dec. 10, 2019; Non-Provisional patent application Ser. No. 17/100,908, filed on Nov. 22, 2020; and the specifications and claims (if any) thereof are incorporated herein by reference.

BACKGROUND OF THE INVENTION Field of the Invention (Technical Field)

Embodiments of the present invention relate generally to improved efficiency for lifting and lowering unbalanced loads.

DESCRIPTION OF RELATED ART

Lifting and lowering of loads has often been facilitated with the use of counterweight (counterbalance) to offset the load, in a manner to reduce the required force to raise and lower the load with the counterweight to be in some state of balance. Whether as in the intentionally unbalanced state, for example, in the Trebuchet beam, a fulcrum machine where a counterweight heavier than the load causes a beam with a fulcrum point to hurl a missile projectile from the opposite lighter beam end when the much heavier counterweight end drops; or in intentionally balanced modes, for example, an elevator, or a beam well pumping unit, often referred to as a “pump jack”, the term “net force” or other synonyms can be used to describe a quantity of positive or negative force required to raise or lower a load after factoring in an attempt to balance or unbalance with counterweight in order to lighten or increase the load. “Gross torque” and other synonyms can be used to describe a quantity of torque required to raise or lower a load without or before an attempt to balance or unbalance with a counterweight—for example, a weight lifting exercise machine whose very purpose is to be heavy.

Gravity is the natural force being countered with the machine's counterbalance force, so with a fixed amount of load and fixed amount of counterweight the machine's required force is relatively constant. Some designs have attempted to improve lifting efficiency in various ways: by varying the angles of pull in the pulling machine, varying the length of linkages in the pulling machine, varying the size of pulleys in the pulling machine, and/or varying the speed reduction of pull in the pulling machine. In the case of beam pumping units which raise and lower a more or less vertical load there is a tipping (fulcrum) point and counterweight effort and load is intended to be in a close state of balance.

Machines designed to do heavy lifting are big and expensive and repairs on worn parts are expensive. The less force that is needed to accomplish the desired work, the smaller the machine components can be, and the less energy can be consumed accomplishing the work, and the less wear and tear on the machine occurs, and all this results in less expense to operate the machine, so designers have tried force-reducing designs in order to improve the economics of the lifting work.

Now we describe some design attempts to reduce the required lifting forces that are variations of both adjustable crank weight and beam weight “conventional” center tipping (fulcrum point) class 1 lever geometry and class 3 lever geometry (rear tipping-fulcrum point) that have attempted to reduce required counterweight in beam well pumping which in operation converts rotary motion of the prime mover, speed reducer, and crank arms, to vertical reciprocating motion of the pitman arms connected to the beam in order to facilitate rod pumping. Besides conventional class 1 geometry these variations can be front-mounted with rear fulcrum points as a class 3 lever, as in the first 1920s air balance units which still use air cylinder pressure as counterweight, and Parkersburg's “Monkey Motion” with fourteen bearing points which was entirely beam weighted with no crank arm weights which made the larger size beam weights bulky. However, both these designs allow more constant effective counterbalance than crank weighted with rotary motion counterweights as used in the 1930s “grasshopper” (Mark II) with class 3 rear fulcrum.

Deeper wells required more counterweight so massive units came of age in the early 1970s when the first sales order for the Mark II 1280 for Union Oil well in Farnsworth, Tex., was obtained by E. L. Hudson which started the era of massive crank weight pumping units when the Mark II's inventor Walter Trout instructed his engineer Joe Byrd to further refine the grasshopper design to accept the largest phased crank counterweight unit ever, and so came the first Mark II 1280.

One problem is that in beam pumped wells the lifted weight is about 1.5 times the weight of the lowered weight due to lifting the weight of the fluid plus the buoyant weight of the sucker rods in the pipe when lifting, but the fluid weight is then held by the downhole pump standing valve when lowered making lifting and lowering unbalanced, so in known references, the difference in counterweight required is split on the up stroke and down stroke which leaves significant unresolved net torque due to the unsolved unbalanced downhole condition.

With conventional beam units, massive effective counterweight is achieved with leverage of adjustable crank weight. But purely beam weighted units were built by Parkersburg and Cabot and others because the effective beam weight is direct and is more constant than rotary crank weight.

A phased crank design for conventional beam unit with class 1 lever center fulcrum point was published by George Eyler and Cabot Corporation in 1963. And an advanced geometry design was published by Bob Gault and Bethlehem Supply in 1965.

These design elements require operating the unit in one direction only and mainly address effective counterweight applied to torque factor, which is a crank angle based multiplier from unit geometry that affects torque calculation at the speed reducer, and sometimes is able to reduce torque over “conventional” designs.

But, the air balance design can reverse direction and the gear teeth in the speed reducer are known for long life. This is partly because with easily adjusted air pressure the counterweight balance is easily maintained close to equal on upstroke and downstroke.

In 1984, Sam Gibbs introduced a wave equation that allowed well controllers to shut off pumping units when fluid in the well bore was low. Thus, variable frequency drives were introduced to seek better efficiency by slowing the pumping units or shutting them off when fluid in the well bore was low. This has led to many other intelligent controllers including speed controllers and soft reversing mechanisms.

All the designs mentioned can achieve a fairly limited increase in efficiency but still leave the problem of downhole unbalanced weight between lifting and lowering. So, there's much room for improvement—including the need for much greater efficiency regarding reduction of torque and net torque, in order to achieve longer lasting components, and reduced operating expense, reduced power consumption, longer stroke lengths and smaller speed reducers.

Some Noteworthy Patents:

Pat. # Date Inventor Class 1,895,181 Jan. 23, 1933 W.C.TROUT 2,134,326 Oct. 25, 1938 R.G. DE LA MATER 74-41 2,155,174 Apr. 18, 1939 W.C. TROUT 74-591 2,179,649 May 04, 1939 W.C. TROUT ET AL 74-41 2,210,661 Sep. 06, 1940 J.L. FINCHER/LUFKIN 74-593 2,213,646 Sep. 30, 1940 A.M. BUTCHER 2,232,245 Feb. 25, 1941 R.G. DELAMATER 74-539 2.293,915 Sep. 25, 1942 E.W. PATTERSON 74-589 2,915,919 Dec. 08, 1959 C.C. MITCHELL 74-590 3,310,988 Mar. 28, 1967 R.H. GAULT 74-41 3,406,581 Oct. 22, 1968 G. EYLER 74-41 4,490,094 Dec. 25, 1984 S.G. GIBBS 417/42, 417/22X, 417/53X, 22-24

BRIEF SUMMARY OF EMBODIMENTS OF THE PRESENT INVENTION

Embodiments of the present invention relate to lifting and lowering loads more efficiently and also more economically than known systems. This invention is the continuous movement by mechanical means of a moving crank shaft force point to a desired advantageous position at a desired advantageous moment to achieve improved geometry resulting in a desired low torque factor for a reduced net torque when lifting or lowering an unbalanced load with a beam with a fulcrum and connected to a load and an effort.

In one embodiment, a walking beam well pumping unit, the lifting and lowering of the well load can be caused by the reciprocating motion of a beam tipping on a fulcrum and with an effort force point.

Potentially reduced net torque might allow longer life speed reducers, smaller speed reducers, and longer reciprocating vertical stroke length and these are both economic and performance benefits.

Objects, advantages and novel features, and further scope of applicability of the present invention will be set forth in part in the detailed description to follow, taken in conjunction with the accompanying drawings, and in part will become apparent to those skilled in the art upon examination of the following, or may be learned by practice of the invention. The objects and advantages of the invention may be realized and attained by means of the instrumentalities and combinations particularly pointed out in the appended claims.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

The accompanying drawings, which are incorporated into and form a part of the specification, illustrate one or more embodiments of the present invention and, together with the description, serve to explain the principles of the invention. The drawings are only for the purpose of illustrating one or more preferred embodiments of the invention and are not to be construed as limiting the invention. In the drawings:

FIG. 1, pumping unit with moving crank shaft force point positioning drive is sandwiched by a speed reducer housing and a speed reducer pedestal with crank arm at 0 degrees. Showing optional embodiments of electric, pneumatic, and hydraulic actuation.

FIG. 2, pumping unit with moving crank shaft force point positioning drive is sandwiched by a speed reducer housing and a speed reducer pedestal with crank arm at 90 degrees. Showing optional embodiments of electric, pneumatic, and hydraulic actuation.

FIG. 3, pumping unit with moving crank shaft force point positioning drive is sandwiched by a speed reducer housing and a speed reducer pedestal with crank arm at 180 degrees. Showing optional embodiments of electric, pneumatic, and hydraulic actuation.

FIG. 4, pumping unit with moving crank shaft force point positioning drive is sandwiched by a speed reducer housing and a speed reducer pedestal with crank arm at 270 degrees. Showing optional embodiments of electric, pneumatic, and hydraulic actuation.

FIG. 5, rodless cylinder positioning drive with optional air, gas, and hydraulic line hookup.

FIG. 6, beam air compressor with air lines, air ports, air reserve tank, and optional auxiliary air compressor or hydraulic pump.

FIG. 7, motorized screw bolt positioning drive.

FIG. 8, rack and pinion positioning drive.

FIG. 9, linear motor positioning drive.

FIG. 10, pumping unit with beam air compressor with moving counter weight assist.

DESCRIPTIVE KEY

-   -   1 torque (force)     -   2 positioning drive     -   3 reciprocating walking beam (lever)     -   4 fulcrum (tipping point)     -   5 carriage (for rodless piston)     -   6 positioning element     -   7 beam angle sensor (inclinometer connected to system         controller)     -   8 air compressor (or hydraulic pump connected to reserve tank)     -   9 well load     -   10 crank weight load     -   11 well     -   12 memory module     -   13 crank position sensor (magnet-transducer connected to system         controller)     -   14 system logic controller     -   15 display     -   16 double acting rodless cylinder (drive positioning assembly)     -   17 ammeter (connected to prime mover)     -   18 crank arm     -   19 crank arm weight     -   20 crank pin (of crank arm)     -   21 moving crank shaft force point (of speed reducer)     -   22 motor sensor (connected to system controller)     -   23 piston foot (mounted on runner)     -   24 cylinder bearing     -   25 effort force point     -   26 piston     -   27 load cell (sensor connected to system controller)     -   28 VFD (variable frequency drive connected to prime mover)     -   29 prime mover (connected to VFD and ammeter)     -   30 system controller (connected to sensors/positioning drive)     -   31 beam air compressor (with air hose to air reserve tank)     -   32 air reserve tank (with air hose to rodless cylinder and beam         air compressor)     -   33 pressure port     -   34 air line or gas line or water line     -   35 piston rod     -   36 horse head     -   37 beam pumping unit (conventional class 1 lever showing         invention)     -   38 samson post     -   39 runner     -   40 speed reducer pedestal     -   41 speed reducer     -   42 electric motor     -   43 enclosure mount     -   44 coupling     -   45 bearing     -   46 screw     -   47 slide     -   48 table     -   49 sensor     -   50 v-belts (from prime mover to unit sheave)     -   51 rack     -   52 pinion     -   53 motor pulley     -   54 pinion pulley     -   55 casing back pressure regulator     -   56 gas line in     -   57 gas return line     -   58 gas sales meter     -   59 pitman arm     -   60 controller connection     -   61 gas liquid scrubber accumulator     -   62 water line     -   63 hydraulic accumulator     -   64 moving beam counter weight     -   65 linear motor electric control     -   66 linear motor positioning drive

DETAILED DESCRIPTION OF THE INVENTION

As used throughout this application, the term “counterbalance” is intended to mean the amount of effective weight the dead weight of the block of steel called “counterweight” 10 must exert to effect a desired result on a well load 9. The term “weight” and “dead weight” when referring to a “counterweight” 10, is used for the sake of simplicity and is not intended to limit the “counterweight” 10, instead, the term “weight” and “dead weight” when used in the context of the “counterweight” 10 is intended to include any and all manners of a “counterweight” 10, including but not limited to reciprocating counterweight, counter weight and counter-weight.

As used throughout this application, the term “net torque” is intended to mean the amount of torque that speed reducer 41 or prime mover 29 must exert to effect a desired result on a well load 9.

As used throughout this application, the term “torque factor” is defined in API Specification 11E, Appendix C, “The torque factor for any given crank angle is the factor that, when multiplied by the load 9 in pounds at the polished rod, gives the torque 1 in inch-pounds at the crankshaft 21 of the pumping unit speed reducer 41.”

As used throughout this application, the term “Permissible Load” is the polished rod load 9 necessary to give a resultant net torque 1 equal to the API rating of the speed reducer 41 with a certain amount of counterbalance 10. This load should be calculated for each 15 degree crank position.

As used throughout this application, the term “unbalanced load” on a beam pumping unit 37 is intended to mean where the load 9 in the lift direction exceeds the load 9 in the return direction.

In accordance with embodiments of the invention, the best mode is presented in terms of the described embodiments, herein depicted within FIG. 1 through FIG. 12. However, the disclosure is not limited to the described embodiments and, upon studying the instant application, a person skilled in the art will appreciate that many other embodiments are possible without deviating from the basic concept of the disclosure and that any such work around will also fall under its scope. It is envisioned that other styles and configurations can be easily incorporated into the teachings of the present disclosure, and only certain configurations have been shown and described for purposes of clarity and disclosure and not by way of limitation of scope.

It can be appreciated that, although such terms as first, second, etc. may be used herein to describe various elements, these elements should not be limited by these terms. These terms are only used to distinguish one (1) element from another element. Thus, a first element discussed below could be termed a second element without departing from the scope of the present invention. In addition, as used herein, the singular forms “a”, “an” and “the” are intended to include the plural forms as well, unless the context clearly indicates otherwise. It also will be understood that, as used herein, the term “comprising” or “comprises” is open-ended, and includes one (1) or more stated elements, steps or functions without precluding one (1) or more unstated elements, steps or functions. Relative terms such as “front” or “rear” or “left” or “right” or “top” or “bottom” or “below” or “above” or “upper” or “lower” or “horizontal” or “vertical” may be used herein to describe a relationship of one (1) element, feature or region to another element, feature or region as illustrated in the figures. It should be understood that these terms are intended to encompass different orientations of the device in addition to the orientation depicted in the figures. It should also be understood that when an element is referred to as being “connected” to another element, it can be directly connected to the other element or intervening elements may be present. It should also be understood that the sizes and relative orientations of the illustrated elements are not shown to scale, and in some instances they have been exaggerated for purposes of explanation.

Embodiments of the present invention can be used in conjunction with the four bar mechanism and linked to work as a reciprocating rod pump.

We are describing an invention which is an apparatus used to reciprocate a downhole rod pump connected to the surface by a rod or rods for the purpose of lifting fluid from a well bore. This invention is an apparatus which is uniquely comprised of a moving crank shaft force point 21.

Of the many applications that embodiments of the present invention apply to, now consider an embodiment of the present invention as applied to class 1 lever, in this particular example conventional crank weight walking beam 3 pumping unit 37 where circular motion is transferred from prime mover 29 to speed reducer 41 moving crank shaft force point 21 and rotating crank arm 18 and then converted to linear motion with crank arm 18 and crank pin 20 articulated with effort force point 25 to walking beam 3, and with this teaching speed reducer 41 net torque and prime mover 29 net torque is reduced by a moving crank shaft force point 21 lowering the torque factor on both adjustable crank weight beam pumping units 37 and beam weight only units.

This device reduces the current problem of high net torque needed to lift and lower the unsolved unbalanced well load 9 in the current practice, by teaching a position-changing moving crank shaft force point 21 whose structurally determined and timed position in reference to the crank pin 20 either forward or aft of the speed reducer moving crank shaft force point 21 being substantially vertical with the effort force point 25 reduces lifting or lowering net torque 1 when the walking beam 3 pumping unit 37 crank arms 18 rotate crank pin 20 and effort force point 25 maintain a vertical relationship.

For beam counterweighted walking beam 3 pumping units 37, load prediction calculations are directly proportional to the effective counterbalance. And calculations for rotary counterweight pumping units can include the API 11E standard equation for calculating net speed reducer 41 torque which is: TN=TF(W−B)−M SIN θ

Where;

θ=Angle of crank arm 18 rotation in a clockwise direction viewed with the wellhead to the right and with zero degrees occurring at 12 o'clock degrees,

TF=torque factor for a given crank angle (from manufacturer's tables or computed from geometric measurements),

B=structural unbalance (from manufacturer or measured),

Tn=Net torque, inch-pounds, at the crankshaft for a given crank angle θ,

W=polished rod load at any specific crank angle θ,

M=maximum moment of the rotary counterweights (from manufacturer or computed from measurements), With these input values Tn=net torque are computed.

The rotational motion of crank arm 18 causes a maximum moment of rotary crank arm 18 weight, crank shaft, and crank pin 20 about the crankshaft whose standard nomenclature is written in thousands of inch-pounds. That maximum moment is nominally the position of the maximum effective crank arm 18 counterbalance at a little less than FIG. 2 at 90 degrees and a little less than FIG. 4 at 270 degrees. FIG. 2 at 90 degrees and FIG. 4 at 270 degrees is nominally the position of maximum net torque and maximum requirement for counterbalance effect. So when the effort force point 25 remains vertically oriented with the crank pin 20 it is decreasing the torque factor to compensate and offset the maximum net torque requirement in the horizontal crank arm 18 position.

A moving crank shaft force point 21 can be retrofitted and installed on already existing units on the existing well installations by using attaching methods such as but not limited to bolts, rivets, weld, and other suitable methods.

And the moving crank shaft force point 21 is desirable to be incorporated in original equipment manufacturing, OEM, on newly manufactured walking beam 3 pumping units 37. Both retrofitted and OEM can employ user discretion to fit the particular specific operational design parameters.

Retrofitted and OEM walking beam 3 pumping units 37 utilizing this invention can potentially allow for longer beam pump stroke lengths and smaller torque capacity speed reducers 13 than those of current practice in known systems because of the increased efficiency with moving crank shaft force point 21 position effectively reducing the required lifting and lowering net torque.

And also with those longer strokes walking beam 3 pump can operate at slower strokes per minute, and also allow the use of reduced prime mover 29 horsepower, so new beam pumping unit 37 designs will want to accommodate the benefits of this invention, where: Load×Distance from tipping point=Counterweight Mass×Distance from tipping point and is called load moment. Current practice rule of thumb ECB (effective counterbalance)˜Bouyant weight of rods+½ fluid load on pump plunger.

Lowest speed reducer 41 torque loads on walking beam 3 pumping units 37 occur at top and bottom of stroke, 0 degrees and 180 degrees, because of low torque factor from unit geometry. And nominal peak speed reducer torque loads occur at high torque factor at about 90 degrees in FIG. 2 and about 270 degree in FIG. 4 crank arm 18 angles which values are desired to be equal when the walking beam 3 pump is balanced in the field at the well using current practice in known systems.

Intentional negative torque 1 caused by deliberate unbalance using a moving counter weight 64 is electrically regenerative and can assist a beam air compressor 31 and negative unbalance may occur when intentionally reducing torque 1, but negative unbalance above speed reducer 41 torque 1 rating can reach diminishing benefits so the recommended control parameters will limit negative torque 1 to be within speed reducer 41 torque rating. Subsequent operating manuals can address details of these and other operational aspects, where: Net torque (Tn)=9.53×kilowatt (kw)×efficiency (eff)/strokes per minute (SPM)×speed variation of power transmission (SV).

-   -   Torque factor (TF) is used to convert polished rod load to         torque (Nm).         Torque due to net well load (TWN)=torque factor (TF)×well load         (WN).         Net well load (WN)=well load (W)−unit unbalance (SU).

FIG. 1 to FIG. 4 and FIG. 10 illustrate an embodiment wherein well 11 is pumped by beam 3, which lifts load 9, which in this particular example is about 1.5 times greater when lifting than that of load 9 when it is being lowered. This is due to lifting the weight of the fluid plus the buoyant weight of the sucker rods in the pipe when lifting up, but that weight is then held by the well tubing in the downhole pump standing valve when being lowered. Thus, in known systems, the difference in load 9 is more or less split on the up stroke and down stroke which leaves a state of significant net torque 1 on prime mover 29 and net gear torque in in the speed reducer 41, due to the remaining unsolved unbalanced condition. Embodiments of the present invention reduce the problem of these high net torque 1 needed to lift and lower load 9 and crank weight 19 at effort force point 25 with a moving crank shaft force point 21 whose moving position maintains a substantially vertical relationship with effort force point 25 reducing torque factor and thus lifting and lowering net torque 1. In one embodiment, prime mover 29 can include but is not limited to an electric motor, an internal combustion engine, a hydraulic motor, combinations thereof and the like. Most preferably, the effort force point 25 is positioned substantially vertical to the crank pin 20. In this embodiment, moving crank shaft force point 21 intelligently changes position so the crank pin 20 maintains positioning substantially vertical with effort force point 25 where it can best cause the most reduction in net torque 1 that is required to lift and lower load 9 and/or crank weight load 10 at moving crank shaft force point 21. FIG. 1 to FIG. 10, in this embodiment, illustrate a class 1 lever having beam 3 that pumps with adjustable crank weights 19 and moving crank shaft force point 21.

FIG. 1 to FIG. 10 illustrate that of the many applications that embodiments of the present invention can apply to, we now consider an embodiment of the present invention as applied to class 1 lever—for example conventional crank weight pumping unit 37 as applied to a reciprocating beam pump, where circular motion is transferred from prime mover 29 to speed reducer 41 and then converted to linear motion with a pitman that is connected from crank pin 20 to effort force point 25 of beam 3 and net torque 1 is reduced by positioning moving crank shaft force point 21. Standard beam fixed weighted units in this description will not be drawn separately because they operate similar to conventional crank weight pumping unit 37. However an embodiment with a moving beam counterweight 64 that assists a beam air compressor is illustrated in FIG. 10.

FIG. 3 illustrates an embodiment that when crank arms 18 are straight down at 180 degrees, which is the position of low torque factor, and moving crank shaft force point 21 is neutrally positioned near the middle of it's range and wherein front end, i.e. the end nearest well 11 of beam 3 is as high as it will go at or about 180 degrees crank angle and is about to re-start the cycle of reciprocating downward. At that moment, moving crank shaft force point 21 begins moving away from near the middle of it's range and is timed to maintain near verticality with the crank pin and arrive at the end of it's range when crank arm 18 reaches near horizontal at either FIG. 2 at 90 degrees or FIG. 4 at 270 degrees to achieve maximum effective offset to the high torque factor in helping to rotate the crank weight. Then, crank arms 18 continue rotating and moving crank shaft force point 21 starts moving back to neutral near the middle of it's range, where it is preferably timed to arrive in neutral near the middle of it's range when crank arms 18 are straight up in FIG. 1 at 0 degrees or in FIG. 3 straight down at 180 degrees.

One method to calculate timing with position selection apparatus comprised of at least one of a vector logic circuit and a moving crank shaft force point 21 position circuit is: Distance from middle of moving crank shaft range to front moving crank shaft force point position/seconds elapse between 90 degree crank positions=feet per second (fps) moving crank shaft force point 25 speed.

Distance from middle of moving crank shaft range of motion to rear moving crank shaft force point 21 position/seconds elapse between 90 degree crank positions=feet per second moving crank shaft force point 21 travel speed Example: 10′/2 seconds=5 feet per second (fps) moving crank shaft force point 21 travel speed

FIG. 1 illustrates that continuing the cycle when crank arms 18 are straight up at 0 degrees, which is the position of low torque factor, moving crank shaft force point 21 is preferably neutrally positioned near the middle of it's range of motion and the rear end of beam 3 is as high as it will go and so will re-start the cycle of reciprocating downward. At that moment, moving crank shaft force point 21 preferably begins moving away from neutral near the middle of it's range of motion and begins moving toward the end of it's range of motion and is preferably timed to maintain near verticality with the crank pin 20 and arrive at the end of it's range of motion when crank arm 18 reaches near horizontal either in FIG. 2 at 90 degrees or in FIG. 4 at 270 degrees to achieve maximum counterbalance effect to offset high torque factor in helping to rotate the crank weight 10. Then, crank arms 18 continue rotating and moving crank shaft force point 21 starts sliding back to neutral, near the middle of it's range of motion, where it is timed to arrive when the crank arms are straight down in FIG. 3 at 180 degrees in or in FIG. 1 at 0 degrees.

Embodiments of the moving crank shaft force point 21 positioning drive 2 can include, but are not limited to FIG. 5 showing a double acting rodless cylinder 16 for low cost with precise control of lightish loads; FIG. 7 a motorized lead screw 46 bolt or ball screw 46 for moderate loads; FIG. 8 a rack 51 and pinion 52 gear drive for heavy loads; FIG. 9 a linear motor positioning drive 66 for control and speed; also two opposing single acting rodless cylinders; tandem double-acting pneumatic cylinders for extended length; double-acting hydraulic cylinder with a hydraulic pump and return reservoir; also magnetic field propulsion; electro magnetism; an electric motor, a reciprocating electric motor; linear motor 66 or servo motor with encoder; cable pullers, chain pullers, and/or belt pullers, with a series of pulleys configured to actuate with beam 3; and other mechanical means consisting of gears, cables, chains, belts, and electric or magnetic drive.

FIG. 5, in one embodiment the movement of moving crank shaft force point 21 is preferably caused by a double-acting pneumatic rodless cylinder in a cylinder positioning drive 16 when the weight is sufficiently low to enable pneumatic actuation. A moving beam counterweight allows use of crank arms 18 without crank arm weight 19 which is lighter. Sensors on cylinder positioning drive 16 can be magnetic to achieve spot positioning as cylinder passes or transducer type sensors for more continuous positioning signals. Pressure ports 33 with nomenclature like 4/2 and 5/2 on cylinder positioning drive 16 open and close as calculated and are instructed to actuate a cylinder of positioning drive 16 to move moving crank shaft force point 21 into position as computed by programmable system logic controller 14. For example, the moment the rodless positioning element 6 passes by a magnetic sensor on the rodless piston carriage 5 a signal indicating that spot position can be sent to the programmable system logic controller 14. Or reaching the end of the range of motion can be sensed and it can then be programmed to reverse and return. Or adding transducers on the rodless piston carriage 5 can detect and measure variations in current and/or voltage which can be converted to signals that indicate the real time rodless pneumatic piston position and speed in the carriage 5. Also speed control muffler, quick exhaust valves, needle valve and flow control fittings can control pneumatic speed. And potentiometer, hall effect sensor, motor controller, lead switch, and limit switch can be used for linear actuator and servo motor position control.

Moving crank shaft force point 21 movement can be controlled by using basic reversal controls coordinated with beam position. In one embodiment, an apparatus and/or system to monitor movement and achieve position control of moving crank shaft force point 21 moving from at or near one or both ends of range of motion can be accomplished by obtaining readings of power use by prime mover 29 by ammeter 17 sent to programmed system logic controller 14 to send signals in FIG. 5 to pressure ports 33 on a rodless cylinder of cylinder positioning drive 16 to which moving crank shaft force point 21 is connected, for positioning moving crank shaft force point 21 to maintain ammeter 17 reading nearest to a low amperage reading throughout the complete reciprocating cycle.

In a redundant configuration, an optimum position of moving crank shaft force point 21 on beam 3 to achieve a reduction in torque factor and net torque 1 can be computed by programmed system logic controller 14 interpreting input data which can include from load cell 27 communicably coupled to load 9, an inclinometer 7 sensing angle of beam 3 and pitman arm 59, transducers 22 and magnets on prime mover 29, sensor 13 on the pumping unit structure sensing crank angle and strokes per minute, and/or signals from magnets or transducers on positioning drive 2. The signals from one or more sensors and load cell 27 are preferably interpreted by programmed system logic in controller 14 to compute optimum lowest value of at least one of the power required of a prime mover and the amount of net gear torque in in the speed reducer 41 by positioning moving crank shaft force point 21, most preferably based on information from measurements, which can include, but is not limited to calculations using wave equation and inserted into a machine learning algorithm program.

One embodiment of the present invention uses programmable system logic controller 14 that processes the previously described measurements and provides them to a sending unit so that the machine learning algorithm communicates instructions to the drive mechanism to control the position of moving crank shaft force point 21 continuously. In this embodiment, programmable system logic controller 14 can, for example, be programmed with the machine learning algorithm such that it will continually process new readings, parameters, and measurements and continually, in real time or near real time, send positioning instructions to positioning drive 2 to position moving crank shaft force point 21 in order to reduce at least one of the power required of a prime mover and the amount of net gear torque 1 in in the speed reducer 41. These measurements can be based on the reduced effective amount of load 9 and effective crank weight load at moving crank shaft force point 21. In one embodiment, a series of downhole measurements, including but not limited to those from load cell 27 on well 11, can be input into programmable system logic controller 14. Another input can optionally include a position of crank arm 18, which can optionally be obtained, at least in part, based on measurements from sensors 13, which can include magnetic transducers, on the pumping unit structure. Other inputs can optionally include the position of beam 3 and pitman arm 59 based on inclinometer 7; ampere measurements from ammeter 17, for embodiments wherein prime mover 29 comprises an electric motor; from vacuum readings for embodiments wherein prime mover 29 comprises an internal combustion engine; and/or one or more measurements stored in memory module 12 for programmable logic controller 14. In one embodiment, display 15 is preferably provided and can be operatively connected to memory module 12 and/or programmable system logic controller 14 for displaying a series of measurements stored in memory module 12 for system logic controller 14, and/or for displaying indicia of one or more values from any other sensor or combinations of sensors used.

In one embodiment, machine learning algorithm can be processed using system controller 30. System controller 30 preferably sends positioning instructions to positioning drive 2 most preferably in real time or near real time. Processed instructions to positioning drive 2 selectively positions moving crank shaft force point 21 so that the position of moving crank shaft force point 21 causes reduced torque factor and net torque 1 as the result of reduced effective load 9 and then also reduced effective crank weight load 10 at moving crank shaft force point 21. System controller 30 preferably includes a programmable system logic controller 14 that is most preferably programmed with a machine learning algorithm and allows the continuous processing of new readings, parameters, and measurements. System controller 30 also preferably includes a sending unit that communicates the processed data to a positioning selection mechanism that is preferably communicably coupled to positioning drive 2, which positions moving crank shaft force point 21 in accordance with the instructions to maintain verticality with the crank pin 20.

In one embodiment, the machine learning algorithm can be processed by system controller 30 with an input for a series of downhole measurements. These inputs can include but are not limited to inputs from load cell 27 on well 11; crank arm 18 position measurements—for example from magnets with sensors 13 mounted on the pumping unit structure; position of beam 3 from inclinometer 7; ampere measurements from ammeter 17, for embodiments wherein prime mover 29 is an electric motor, and/or vacuum readings for embodiments wherein prime mover 29 is an internal combustion engine.

In one embodiment, the machine learning algorithm can be processed by a measurement input that can store a series of measurements in memory module 12 for programmable system logic controller 14.

In one embodiment, system controller 30 can process the series of measurements stored in memory module 12 for system logic controller 14 with a machine learning algorithm such that display 15 is operatively connected to the system controller with memory module 12 and programmable system logic controller 14.

The rotational motion of crank arms 18 cause a maximum moment of rotary crank weights 19, crank shaft 21 and crank pins 20 about moving crank shaft force point 21 in inch-pounds, which is nominally the maximum effective moving crank shaft force point 21 counterbalance in FIG. 2 at about 90 degrees or in FIG. 4 at about 270 degrees. And in FIG. 2 at about 90 degrees or in FIG. 4 at about 270 degrees is nominally the position of maximum torque and effective counterbalance. So when the programmed logic commands the position of moving crank shaft force point 21 to be positioned to maintain the verticality of the crank pin 20 with effort force point 25, it is lowering torque factor and increasing effective crank counterweight 10 to offset the maximum amount of net torque 1 in the horizontal position to raise load 9, and, vice versa when commanded to maintain verticality of the crank pin 20 with effort force point 25 to raise the effective crank weight 10.

FIG. 5, in one embodiment, system positioning drive 16 preferably comprises a rodless cylinder, which preferably sits on top of speed reducer pedestal 40 or runner 39.

Moving crank shaft force point 21 is a member of a speed reducer 41 which is preferably disposed above positioning drive 16 with a cushion on each end to soften reversals. In one embodiment, movable positioning element 6 of positioning drive 16 can comprise rodless positioning element 6, which can be magnetic and thereby securely attract metallic moving crank shaft force point 21 to follow its movements.

FIG. 5, in one embodiment air pressure for double-acting rodless cylinder of positioning drive 16 can be supplied by any capable and/or suitable air supply with an air line 34 to reserve tank 32 and an air line 34 to a valve such as 4/2 or 5/2 valve. In a well 11 pumping embodiment when the beam pumping unit 37 is pumping a well 11 that produces some natural gas where the natural gas is normally separated from fluid in the casing, the gas can be used at flowline pressure, nominally 30-40 psi, or can be caused to accumulate hundreds of pounds of pressure via a casing back pressure regulator 55 and then pass through a gas liquid scrubber 61 as can be a standard operating procedure in special production scenarios (for instance in some cases to reduce sand production, decrease oil viscosity, decrease BS&W, to push casing fluid level lower, etc.), and then routed under pressure through a gas line in 56 to actuate a double-acting rodless cylinder type positioning drive 16, and then the higher pressure gas is vented back to a lower pressure gas return line 57, then routed through the natural gas sales meter 58 and into the gas sales line. This method has efficiency because it uses the already existing work that produces gas under pressure by the pumping system to energize the actuator without requiring additional energizing work.

FIG. 5, in one embodiment hydraulic pressure to actuate the double-acting rodless cylinder type positioning drive 16 can be accumulated from pumping a fresh water well with a hydraulic accumulator 63 and back pressure. This method has efficiency because it uses the already existing work that produces water under pressure by the pumping system to energize the actuator without requiring additional energizing work.

FIG. 6, in one embodiment, compressed air is preferably provided by an air compressor 8, which can optionally include a piston type beam air compressor 31 powered by beam 3. In one embodiment, a double-acting single piston-cylinder can be used as beam air compressor 31, and is preferably pivotably connected at its top by a cylinder bearing 24 to beam 3 and the lower extremity is preferably pivotably anchored with a piston foot bearing 23 connected to the beam pumping unit 37 structure runner 39. With this configuration, air can be compressed by beam 3 actuated piston air compressor 31 using power from beam 3 which actuates the movable positioning element 6 the length of the positioning drive 16. A two stage beam air compressor 31 embodiment can be used to increase pressure. When there is intentional negative torque achieved with a moving beam counterweight 64 this method has efficiency because it can use the already existing work by the pumping system to reciprocate the beam to energize the beam air compressor 31 to buck pressure without requiring auxiliary power.

Of course other configurations of air compressors can be used and will provide desirable results, including but not limited to configurations in which the air compressor is connected by one or more belts to the pumping unit sheave being powered by prime mover 29 or to the sheave of prime mover 29 itself. Of course, motive power source 8, which can optionally be a compressed natural gas source and including but not limited to an air compressor or hydraulic pump, can be provided independent of prime mover 29. In one embodiment, the air compressor can include one or more rotary screw and/or reciprocating air compressors.

For embodiments that employ a double-acting hydraulic cylinder with a two-direction valve, the hydraulic pressure can be supplied from motive power source 8, which can optionally be a beam actuated piston hydraulic compressor 31 hydraulic pump, most preferably with a pressure relief and return reservoir 32.

FIG. 7 illustrates an embodiment when positioning element 6 of positioning drive 2 can comprise nut of screw 46 rotating with bearings 45 with coupling 44 to an electric motor 42, for instance a servo motor or stepper motor and encoder, held with a mount 43. Moving crank shaft force point 21 with a position sensor 49 can be connected to a table 48 sandwiched between moving crank shaft force point 21 and moving positioning element 6 on slide 47.

FIG. 8 illustrates an embodiment when positioning element 6 of positioning drive 2 can comprise rack 51 on guides 47 and pinion 52 with an electric motor 42, for instance a servo or stepper motor and encoder, which can be connected to a table 48 sandwiched between moving crank shaft force point 21 and rack 51. Electric motor 42 can be a direct gear drive or use belt and motor pulley 53 and pinion pulley 54.

Moving crank shaft force point 21 can be connected to the table 48 sandwiched between moving crank shaft force point 21 and positioning element 6 by bolt, rivet, weld, magnetism or other preferably removable fastener.

In one embodiment, the position of moving crank shaft force point 21 is preferably precisely correlated and timed with the strokes per minute of beam 3 pump. Variable frequency drive 28 can slow the upward or downward speed of load 9 in mid stroke when moving crank shaft force point 21 position causes reduced net torque 1. Variable frequency drive 28 with programmable system logic controller 14, which is preferably programmed for this application, can include a component of the preferred configuration. This is because desired strokes per minute of beam 3 can be controlled in coordination with data from load cell 27 via the chosen speed of prime mover 29. This configuration increases well management and is available when variable frequency drive 28 is integrated with a sensor package. This embodiment can possibly allow a longer pump stroke without increasing torque 1 in the speed reducer 41 and prime mover 29 and can possibly allow a slower pumping speed which is sometimes operationally desirable.

This moving crank shaft force point 21 can be installed on already existing pump jacks at existing well installations by attaching positioning drive 16 to either the existing pedestal 40 or runners 39, or adapting any other workable method to move moving crank shaft force point 21. Alternatively, moving crank shaft force point 21 can be part of original equipment on newly manufactured reciprocating beam pumping units 37. This may allow OEM design to have longer beam pump stroke lengths and smaller net torque 1 capacity speed reducers 41 than those prior to this invention. This is because of the increased efficiency with the moving crank shaft force point 21 position effectively reducing the lifting and lowering torque factor and net torque 1. And, with those longer strokes possible, the pump jack can operate at slower strokes per minute, possibly reducing tubing and rod wear, and also allow the use of lower horsepower prime mover 29. As such, reciprocating beam pumping unit 37 design can possibly be improved with longer stroke lengths and smaller speed reducers 41, to accommodate the benefits of embodiments of this invention.

Lowest speed reducer 41 net torque 1 loads on reciprocating beam 3 pumping units 37 occur at top and bottom of stroke, in FIG. 1 at 0 degrees and in FIG. 3 180 degrees, because of low torque factor from unit geometry. And with the current art nominal peak net torque 1 loads on speed reducer 41 occur at high torque factor in FIG. 2 at 90 degrees crank arm 18 angles and in FIG. 4 at 270 degrees crank arm 18 angles which are substantially equally high torque values when reciprocal beam 3 pump is operated normally in the current art. Embodiments of the present invention can achieve reduced net torque 1 in FIG. 2 at 90 degrees crank arm 18 angles and in FIG. 4 at 270 degrees crank arm 18 angles because of lower torque factor.

-   -   The following equations further describe an embodiment of the         present invention:         Net torque (Tn)=9.53×kilowatt (kw)×efficiency (eff)/strokes per         minute (SPM)×speed variation of power transmission (SV).     -   Torque factor (TF) is used to convert polished rod load to         torque (Nm).         Torque due to net load (TWN)=torque factor (TF)×load (N).         Net load (N)=load (W)−unit unbalance (SU).     -   In one embodiment, consider:         Pressure (P)=Force (F)/Area (A).         Force (F)=Pressure (P)×Area (A).         Load×Distance from tipping point=Counterweight Mass×Distance         from tipping point and is called load moment.     -   A “Rule of thumb” for top of the head calculation in the field:     -   ECB (effective counterbalance) Buoyant weight of rods+½ fluid         load on pump plunger.

The “Permissible Load” is the polished rod load necessary to give a resultant net torque equal to the API rating of the reducer with a certain amount of counterbalance. This load should be calculated for each 15 degree crank position. The formula used for this calculation is

${W\left( {{Permissible}\mspace{14mu}{Load}} \right)} = \frac{{{Net}\mspace{14mu}{Torque}} - {{M\;{Cos}} \ominus}}{TF}$

-   -   where:

Net Torque=API rating of the gear reducer

M=Counterbalance torque at 90° (FIG. 2) or 270° (FIG. 4) (Determined by counterbalance requirements for a particular pumping application)

θ=Crank Angle

W=Permissible Load (Polished rod load required to give net torque equal to rating of the gear reducer)

Net torque for a pumping unit is calculated by the following API formula: Net Torque=(TF)(W)−(M Cos θ) where:

TF=Torque Factor

W=Polished Rod Load

M=Counterbalance Torque at 90° (FIG. 2) or 270° FIG. 4)

θ=Angle of crank, starting with 0° at vertical position and reading clockwise.

Net torque can be found by this formula when polished rod load, counterbalance torque and crank angle are known. The formula is used to find net torque from dynamometer cards.

The foregoing embodiments have been presented for the purposes of illustration and description. They are not intended to be exhaustive or to limit the invention and method of use to the precise forms disclosed. The embodiments have been chosen and described in order to best explain the principles and practical application in accordance with the invention to enable those skilled in the art to best utilize the various embodiments with expected modifications as are suited to the particular use contemplated. The present application includes such modifications and is limited only by the scope of the claims. Although the foregoing discussion describes the most preferred locations of moving crank shaft force point 21 at various times in the pumping cycle, it is important to understand that such preferred locations are merely described for illustration purposes and desirable results can be achieved when moving crank shaft force point 21 is in approximately such locations with respect to effort force point 25.

Although the invention has been described in detail with particular reference to these preferred embodiments, other embodiments can achieve the same results. Variations and modifications of the present invention will be obvious to those skilled in the art and it is intended to cover in the appended claims all such modifications and equivalents. The entire disclosures of all references, applications, patents, and publications cited above are hereby incorporated by reference. 

What is claimed is:
 1. A moving crank shaft force point on a walking beam pumping unit comprising: a walking beam with an effort force point, a load, a fulcrum; a moving crank shaft force point connected to a rotating crank arm; said crank arm pivotably connected to a rotating crank pin providing a means for a radius with the crank shaft; said crank pin pivotably connected to an effort force point providing a means to reciprocate the load; and a crank shaft force point positioning drive configured to slidably urge the crank shaft force point forward and aft providing a means to dispose said crank pin substantially vertical to the effort force point.
 2. The moving crank shaft force point on the walking beam pumping unit of claim 1, whereby: said moving crank shaft force point further including control for providing a means for a substantially vertical disposition of the crank pin pivotably connected to the effort force point; the moving crank shaft force point connected to the crank arm such that said pivotably connected crank pin is substantially vertically disposed to the effort force point in real time; and said substantially vertical disposition provides efficient geometry thereby achieving a low torque factor.
 3. The moving crank shaft force point on the walking beam pumping unit of claim 1, such that: a substantially vertical disposition of the crank pin pivotably connected to the effort force point provides a means for reduced torque; the reduced torque provides a means for reducing at least one of power required of a prime mover and an amount of net gear torque in a speed reducer; the reduced torque provides a means for reducing required counterweight; and the reduced torque provides a means for increasing permissible load.
 4. The moving crank shaft force point on the walking beam pumping unit of claim 1, whereby said crank shaft force point positioning drive is sandwiched by at least one of a speed reducer housing and a speed reducer pedestal and the speed reducer pedestal and a runner.
 5. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive is a pneumatic effort positioning drive whose actuation comprises: an auxiliary air compressor; a beam air compressor; a moving beam counterweight providing a means to assist said beam air compressor; and compressed gas a well.
 6. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive comprises at least one of: compressed gas from a well providing a means for pneumatic repositioning of the moving crank shaft force point; pressurized fluid from an auxiliary hydraulic pump providing a means for hydraulic repositioning of the moving crank shaft force point; and pressurized fluid pumped from a well providing a means for hydraulic repositioning of the moving crank shaft force point.
 7. The moving crank shaft force point on the walking beam of a pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive comprises a rodless piston.
 8. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive comprises a double-acting rodless piston.
 9. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive comprises a rack and pinion.
 10. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive comprises a lead screw.
 11. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive comprises magnetic field propulsion.
 12. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive comprises an electric motor.
 13. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive comprises at least one of two opposing single acting rodless cylinders; tandem double-acting pneumatic cylinders; ball screw; electro magnetism; encoder; gears; cables; chains; belts.
 14. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 4, whereby said positioning drive comprises: a system controller comprising at least one of: a sensor configured to sense at least one of power required of a prime mover and an amount of net gear torque in a speed reducer, logic programmed to achieve lowest power required of the prime mover and the amount of net gear torque in the speed reducer, and a variable frequency drive.
 15. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 14, whereby: said system controller logic is configured to provide a means for said positioning drive to dispose said moving crank shaft force point such that a downward force substantially vertical to the crank pin exerts on a rear portion of the walking beam during an upstroke of the walking beam pumping unit, the rear-portion of the walking beam being an opposite end-portion of the walking beam from a head of the beam pumping unit.
 16. The moving crank shaft force point on the walking beam pumping unit with the crank shaft force point positioning drive of claim 14, whereby: said system controller logic is configured to provide a means for said positioning drive to dispose said moving crank shaft force point such that an upward force substantially vertical to the crank pin exerts on a rear portion of the walking beam during a down stroke of the walking beam pumping unit, the rear-portion of the walking beam being an opposite end-portion of the walking beam from a head of the beam pumping unit. 